Gas turbine with acoustic surge control



Oct. 1, 1957 A. G. BODINE, JR

GAS TURBINE WITH ACOUSTIC SURGE CONTROL Filed Nov. 25, 1952 INVENTOR.41.5527 61 Boom/E Je.

United States Patent D GAS TURBINE WITH ACOUSTIC SURGE CONTROL Albert G.Bodine, Jr., Van Nuys, Calif.

Application November 25, 1952, Serial No. 322,518

7 Claims. (Cl. 60--39.09)

This invention relates generally to the compressors of gas turbines,particularly, though not limited to, the compressors of aircraftturbo-jet engines. It will be evident, for example, that the inventionis equally applicable to turbines for generating shaft horsepower.

Both centrifugal and axial flow compressors for modern turbo-jets sufferfrom surging, which is an unstable flow condition exhibiting pulsationsof flow which are very detrimental to operation. Ways have been foundfor avoiding the surge region with turbo-jets operating at comparativelylow compression ratios, but with the present design trend toward highercompression ratios, the problem of surging has become most serious, andno satisfactory solution has to my knowledge been heretofore proposed.

The general object of the present invention is accordingly the provisionof means for controlling or suppressing the phenomena of surging inturbo-jet compressors.

The problem of surging has been investigated by others, and certainsignificant facts have been authoritatively reported. For example, R. O.Bullock and H. B. Finger, of Lewis Flight Propulsion Laboratory, NACA,presented a paper at the SAE National Aeronautic Meeting, New York,April 16, 1951, in which was reported their finding that the pulsationsin centrifugal superchargers occur at a frequency which is high forsmall volume, and low for large volume, and their further finding that,in general, the results of the supercharger investigations wereapparently applicable to axial flow compressors. See SAE QuarterlyTransactions, April 1952.

These investigators did not, however, draw the conclusion that theirexperimental data evidenced an acoustic resonant phenomenon, with thecompressor enclosure functioning as a resonant acoustic cavity. Neitherdid they point out that such surging might be controlled byuse ofacoustic remedies for cavity resonance, which is the basic concept ofthe present invention.

The present invention therefore starts with the assumption that acentrifugal or axial flow compressor is capable of behaving as aresonant acoustic cavity. An acoustic cavity may be defined for presentpurposes as a space, chamber or conduit having reflective boundaries ofsuch nature that an acoustic standing wave may be set up therein at oneor more resonant frequencies, or such that it behaves as a Helmholtzresonator. It is known generally that gas driven through a resonantcavity is capable of resonating the cavity. I have found that the flowbecomes unsteady, and surges at a resonant frequency of the cavity, thesurge frequency varying inversely with the size of the cavity. Now, in acompressor, the rate of energy delivery of a compressor vane or blade tothe air stream is a function of air stream velocity. If air streamvelocity pulsates, the energy delivery from blades to air stream mustpulsate accordingly. The back reaction of the air stream on the bladesis thus a fluctuating factor, whose magnitude is proportional to theperiodic deviation of flow velocity from its median value. The energydelivery from the blades accordingly has a fluctuating component, ofperiodicity equal to the acoustic resonant frequency of the compressorcavity. The effect is regenerative, in that the periodic fiow pulsationsare initiated in the first instance by the mere gas flow through thecavity, but, upon inception of periodic reaction with the blades, aregreatly augmented and then maintained at high amplitude by the periodiccharacter of the blade-to-air stream energy delivery. The discovery ofthis phenomenon is a very important part of my invention. The phenomenonis not unlike the drive of a violin string by a bow, where the vibrationfrequency of the string is set by the resonant frequency of the drivenmember (string) while the driver (bow) delivers its energy with afluctuating energy flow, in step with the vibrations of the string. Thestring vibrates at its own resonant frequency as a result of beingdisturbed, the energy delivery from the constantly moving bow to thestring is periodic because of the vibratory motion of the string, andthe amplitude of the vibrations varies with the power of the bow. Insimilar fashion, the air surges at the resonant frequency of the cavityas a result of some disturbance; and because of this surging, the energyflow from blades to air stream is periodic, and the surge amplitudedepends upon the power with which the blades are driven.

Surging, therefore, is a phenomenon initiated by resonantcharacteristics of the compressor cavity, and carried to high amplitudeby the resultant periodic character of the blade-to-airstream energyflow.

Such periodic surging is not only detrimental to the thermodynamiccycle, and the operational characteristics of the turbo-jet as a whole,but is also detrimental to the blades themselves. It has been explainedhow the periodic energy flow from the blades to the air stream resultsin a periodic fluctuation of back reaction on the blades. This periodicfluctuation of reaction may be either at the fundamental resonantfrequency of the compressor cavity, or at some higher harmonic. Theperiodic fluctuation of reaction tends toward vibration of the blades,and particularly if some harmonic frequency should coincide with anatural resonant frequency of the blade, the blade may be set into veryserious vibration. Any vibrational tendency at the blades is of coursehighly undesirable, not only because it would be further conducive toflow instability, but because of blade fatigue.

The present invention provides acoustic means for causing the gasoscillations in the compressor cavity to subside. This is accomplishedby using, in combination with the compressor cavity, a wave or gasvibration attenuator. This may take the form of a resonant absorber,tuned to the natural resonant frequency of the compressor cavity andhaving the ability to destroy or materially suppress the resonantpeaking characteristics of the chamber. In the language of those versedin the acoustic art, my vibration attenuator materially reduces the Q ofthe compressor cavity, the factor Q being understood to denote the ratioof energy stored to energy dissipated per half-cycle of the vibrationwhich is taking place. It does this by dissipating a large amount of thevibrational energy on each half-cycle, and in consequence, the amplitudeof any gas vibration is kept Within harmless limits.

An important object of the invention is to reduce the acoustic Q of thecompressor cavity and thereby prevent or reduce the tendency of the D.C. flow compressor to convert a portion of its energy into A. C. flow.

It is accordingly not the intention merely to dissipate acoustic energyafter it is generated. Rather, it is the intention to break the chain ofmutual cooperation between cavity resonance and unsteady or periodicenergy delivery by the blades, so that the cavity cannot induce theblades 3 to generate sound in the first place. This I accomplish byreducing the acoustic resonance of the cavity;

In the drawings,

Figure 1 is a longitudinal section of a turbo-jet engine embodying theinvention;

FigureZ is an enlarged detailof'Figure l;

Figure 3 is a view of the compressorrotor in the direction of arrows3--3 in Fig. 2; and

Figure4-is an enlargeddetail'section of an attenuator.

In the drawings, numeral designates generally a somewhatdiagrammatically illustrated turbo-jet engine, which is illustrative ofthe invention. This engine has an axial flow compressor 11', a burnersection 12, a bladed turbine rotor-13, and a jet discharge orifice 14,turbine rotor 13 beingonv shaft 15 which carries and drives the bladedrotor 16 of compressor 11'. The compressor and turbine rotors turn as aunit, supported by bearings 17 and 18:

The turbine hasan external shell 20, including a forward cylindricalsection 21 open at the forward end for air intake to compressor 11, anenlarged intermediate section 22 around the burner section, and aconverging rearwardsection 23 terminating in orifice 14.

The compressor rotor is, in general form, of a conven: tional typehaving axially spaced disks 25 shrink-fitted onto shaft 15 and carryingat their peripheries suitably anchored'compressor blades 26'. Thesecompressor blades 26 turn between stationary compressor blades 27suitably supported inside the compressor section 21 of the shell. Theannular air duct between the shell wall and the peripheries of the disks25' converges toward the burner section, in the usual manner, as clearlyillustrated. The spaces between the disks 25, excepting for certainpresently described acoustic attenuator formations, are relativelysmall, e. g., from blade to blade.

The burner section 12 includes an inner annular burner wall 30, formingan annular combustion chamber space or duct 31, and provided with anysuitable means of support (not shown) to the outer shell. Within thisannular space 31 are burners 44 comprising cylindrical tubes closed attheir forward ends and open toward the rear, these being suitablysupported, as by means of webs 45. It will be seen that the airdischarged rearwardly from the compressor enters the forward endof duct31, and that the airflow, including products of combustion from theburner tubes, leaves the rearward end of the duct 31 through stationaryturbine blades 32 to impinge on turbine rotor blades 33; thence flowingrearwardly through the space between-inner tail cone 34 and theconvergent section of the shell, to be finally discharged at orifice 14.

A fuel nozzle discharges fuel into the head end of each burner 44, thefuel being supplied by fuel line 56 from fuel control unit 57 and fuelpump 58. Air enters the burners 44 through intake ports 59. Combustionof the fuel and air mixture within burner 44 results in the discharge ofhigh pressure combustion gases in a rearward direction to drive theturbine rotor, and eventually to be jet-discharged at 14.

Such a turbine is often subject to severe surging, as mentionedhereinabove. The surge frequency is relatively low, typically within therange of 10l000 C. P. S.

My studies have shown that surging in such turbines can take place invarious modes, either within the compressor as a discrete acoutic unit,or within the compressor and combustion sections acoustically coupledtogether, i. e., acting conjointly as a unitary acoustic device.

Considering the first mentioned mode, the shell portion 21 and disk andblade surfaces form an acoustic cavity, open at both ends, and thiscavity possesses, among other probabilities, a fundamental resonantfrequency for a longitudinal half-wave mode of gas oscillation. Such amode is characterized by the presence of a half-wave length standingwave, with velocity antinodes (regions of maximum gas oscillation) atboth ends, and a pressure antinode (region of maximumpressureamplitude.oscillation) at some position intermediate the twoends. Along with" this fundamental, harmonics may develop, givingadditional pressure antinodes along the length of the compressor.

There are two possible principal modes with the compressor and burnersections coupled together or acting as a unitary acoustic device. Thefirst of these is that characteristic of a Helmholtz cavity, where theneck of the cavity is formed by the compressor duct, and the body of thecavity is formed by the somewhat enlarged burner duct. In this case, thegas oscillates back and forth through the compressor duct, and the gasbody within the burner duct experiences a substantialv pressure cycle.The velocity and pressure oscillations occur at the resonant frequencyof the compressor and burner ducts considered as the neck and'body,respectively, of a Helmholtz resonator. Finally, as the second mode withthe compressor and burner sections acting together as a single orcomposite resonant cavity, alow frequency mode canoccur with thecompressor and burner sections forming a long half-wave length conduit,with velocity antinodes at the two ends, and a pressure antinode nearthe junction of the compressor and burner sections.

The invention provides acoustic wave attenuator means in combinationwith these acoustic cavities or ducts formed'by the compressor and bythe compressor coupled with the burner section.

In general,,where a distinct wave pattern is either ascertainable orprobable, the most effective location for such acoustic means is in theregion of pressure antinodes of the gas oscillation modes or patternswhich are to be controlled. Accordingly, for the purpose of controllingthe two modes which involves the compressor and burner sectionsacoustically coupled to one another, I here show, as illustrative. ofthe invention, a plurality of attenuators 60, coupled to the system atthe pressure antinode regions at or near the junction of the burnersection with the compressor section. These attenuators, being fairlyclosely coupled to the compressor section, may also eifect a degree ofcontrol over acousticwave patterns set up in the compressor section as aunit.

In the illustrative embodiment, the attenuators oomprise horns 61,preferably approximately of exponential taper, with attenuativetermination-s 62. The taper is calculated, according to well knowndesign procedure, to accept wave frequencies of the wave or waves whichare to be attenuated, and to convey such waves, without substantialreflection back, to the attenuative termination device 62, where thewaves are dissipated by conversion of their wave energy into heat. Theattenuative termination devices may consist of a packing 63 of some suchfibrous sound wave absorptive material as fiber glass, located withinthe constricted throat of the horn. Preferably, however, .andparticularly for any cases in which the wave amplitudes are sufiicientlyhigh to be destructive of the packing material, I prefer to employ along attenuator tube 64 connected to the throat of the horn, andpreferably coiled, as shown in Fig. 4. This attenuator tube, if providedwith certain rather critical ranges of dimensions, will possess strongattenuation characterist-ics without being of undue or unwieldy length.In general, the requirements are met if the horn has an area ratiobetween mouth and throat of at least 20 to 1, and of a length at leastone-quarter wave length, and the tube has a lateral dimension of about.06 wave length and a length of two wave lengths.

:For principal control'of the sound wave pattern within the compressorsection, I form resonant absorbers between the disks 25 of thecompressor. In the illustrative embodiment, these absorbers 70 are inthe nature of Helmholtz resonators. To form them, opposite faces of thedisks 25 are provided with annular channels or concavities 71. Theseconcavities form the bodies or chambers of the Helmholtz resonantabsorbers 70. The constricted" necks of the absorbers 70'are supplied bythe narrow spaces 72 between the opposed lateral surfaces of the rimportions of the disks. Preferably, to reduce circular motion of thegases between the disks about the central axis, the disks are formed,within the concavities 71, with radial baffies 74. This formationreduces any tendency for setting up of circular modes of gas oscillationwithin the chamber of the Helmholtz resonator.

The dimensions of the Helmholtz resonators so formed are adjusted inaccordance with well known acoustical practices to be resonant orresponsive to the wave frequencies prevalent within the compressorsection. 'If there are more than one wave frequency to be attenuated,some of the Helmholtz resonators can be adjusted to each frequency to becontrolled.

It was previously mentioned that the most effective location for a soundwave attenuator to be coupled to an acoustic cavity or duct forsuppression of a wave pattern therein is the region identified with apressure antinode of the wave pattern. As also previously mentioned, thecharacteristic fundamental sound wave vibration mode or pattern hasvelocity antinodes at the two ends of the compressor duct, and apressure antinode at an intermediate region of that duct. And ifharmonics are present, additional pressure antinodes will appear atadditional locations along the compressor. It will 'be seen that by useof the row of Helmholtz resonant absorbers provided by the invention,extending the full length of the axial flow compressor, there is certainto be one or more of such absorbers in the immediate region of eachpressure antinode within the compressor, even though the preciselocation of such pressure antinodes should not be exactly known.

The Helmholtz resonant absorbers as described function to verymaterially suppress or attenuate the resonant wave frequencies developedwithin the compressor section. In many cases, the vibration modedeveloped within the compressor section as an individual acoustic unitwill be the chief or only offender. In this case, the resonant absorbersemployed within the compressor section will fully suffice to correct theengine.

In cases in which, to any serious extent, other modes are present,particularly modes involving the burner section, additional acousticattenuators, such as those indicated at 62, may be employed.

With the use of acoustic wave attenuators, as described, properly tunedor shaped to be responsive to the offending wave frequencies, thesurging of a gas turbine is forced to subside, and desired stabilitythereby attained.

The present application is a continuation-in-part of my priorapplication entitled Control of Combustion Instability in Jet Engines,filed June 16, 1951; under Serial No. 231,954.

I claim:

1. A gas turbine having an axial flowcornpressor comprising a bladedstator and a bladed rotor between which is an annular axial flow ductwithin which gas surging tends to occur at a predetermined wavefrequency, said rotor comprising an axially disposed shaft and aplurality of axially spaced disks mounted thereon, said disks fitted attheir peripheries with blades for coaction with the blades of thestator, said disks having configurations in opposed side surfacesthereof forming Helmholtz resonators having enlarged chambers andconstricted necks, said necks communicating with said axial flow duct,and said Helmholtz resonators having a resonant response to saidpredetermined wave frequency.

2. A gas turbine having an axial flow compressor comprising a bladedstator and a bladed rotor between which is an annular axial flow ductwithin which gas surging tends to occur at .a predetermined wavefrequency, said rotor comprising an axially disposed shaft and aplurality of axially spaced disks mounted thereon, said disks fitted attheir peripheries with blades for coaction with the blades of thestator, opposed faces of adjacent disks having opposed annular channelssunk therein to form the bodies of Helmholtz resonant absorbers havingan attenuative frequency response for said predetermined frequency, saiddisks having opposed close spaced annular surface portions outside saidchannels to form the constricted necks of said Helmholtz absorbers, andsaid Helmholtz absorbers having a resonant response to saidpredetermined wave frequency.

3. The subject matter of claim 2, including also a plurality ofangularly spaced radially disposed baffles formed on the confrontingsurfaces of the disks within said annular channels.

4. A gas turbine having an air flow pass-age including a bladedcompressor section, in which passage a periodic surge flow phenomenatends to occur at a characteristic sound wave frequency which sound wavefrequency is of low frequency range relative to blade hum and whichsound wave frequency is a function of cavity resonance gas vibrationsreacting upon the compressor blade aerodynamic drive characteristicsunder conditions of high compression and relatively low flow such asoccurs near aerodynamic stall of the blades during high load oraccelerating conditions in ga turbines, and acoustic wave attenuationmeans so constructed and arranged as to have an attenuat-ive frequencyresponse for said sound wave stall surge frequency, which attenuationmeans is acoustically coupled to said air flow passage.

5. The subject matter of claim 4, wherein said attenuation means iscoupled to said air flow passage in the region of said bladed compressorsection.

6. The subject matter of claim 4, wherein said air flow passage has aburner section following said compressor section, and wherein saidattenuation means is coupled to said passage in the region of thejunction between said compressor and burner sections.

7. The subject matter of claim 4, wherein said air flow passage has aburner section following said compressor section, and wherein saidattenuation means is coupled to said passage within said burnersect-ion, but in acoustic coupling relationship to said compressorsection via the junction between said compressor and burner sections.

References Cited in the file of this patent UNITED STATES PATENTS2,171,342 McMahon Aug. 29, 1939 2,225,398 Hamblin Dec. 17, 19402,330,701 Gerber Sept. 28, 1943 2,453,524 McMahon et al. Nov. 9, 19482,543,755 Berger Mar. 6, 1951 2,570,241 Hutchinson Oct. 9, 19512,575,682 Price Nov. 20, 1951 2,579,049 Price Dec. 18, 1951 FOREIGNPATENTS 261,468 Switzerland May 15, 1949

